Change speed gearing and control



June 18, 1940. THOMPSON 2,204,872

CHANGE SPEED GEARING AND CONTROL Filed April 1, 1938 8 Shets-Sheet 1June 18, 1940. E. A. THOMPSON CHANGE SPEED GEARING AND CONTROL 2 aSheets-Sheet 2 Filed April 1, 1938 3nventor Gttornegs June 18, 1940. E.A. THOMPSON CHANGE SPEED GEARING AND CONTROL Filed April 1, 1938 8Sheets-Sheet 3 Zmventor 54 21 QZom uzm w, Gttornegs June 18, 1940. E.THOMPSON CHANGE SPEED GEARING AND CONTROL Filed April 1, 1938 8Sheets-Sheet 4 June 18, 1940.

E. A. THOMPSON CHANGE SPEED GEARING AND CONTROL Filed April 1, 1938 8Sheets-Sheet 5 Snnenfo:

644'! Q Xian 250m 3Q I d 'I I I Gttoxnegs J 18, 1940- E. A. THOMPSON2,204,872

CHANGE SPEED GEARING AND CONTROL Filed April 1, 1938 8 Sheets-Sheet 6 I3nventor June 18,1940. 5 THOMPSON 2,204,872

CHANGE SPEED GEARING AND CONTROL Filed April 1, 1938 8 Sheets-Sheet 7 Q*3; Zinnmtor Gttomegs Patented June 18, 1940 UNITED STATES CHANGE SPEEDGEABING AND CONTROL Earl A. Thompson, Bloomfield Hills, Mich, as-

signor to General Motors Corporation, Detroit, Micln, a corporation ofDelaware Application April 1, 1938, Serial N0.'199,355

50 Claims.

The invention relates to the controls for change speed gearingmechanism, in particular for motor driven vehicles. It is directed toimprovements wherein independently acting automatic and manual controlmembers are coordinated to produce an ideal, fully automatic selectionof speed ratio in step ratio gearing, yielding acceleration or economyaccording to the demand for torque on the part of the operator. It iswell-known that present day motor cars, as an example, are equipped withstep-ratio gearing which requires considerable personal attention on thepart of the driver, if the varying traffic conditions are to be met. Theinvention represents a number of relli lated improvements wherein agraduated automatic selection of speed ratios is always available forany driving condition, with a minimum of shock during the ratiotransitions, and a minimum of attention on the part of the driver.

The invention provides smooth, completely automatic shift among allforward speed ratios, wherein the factor of driver-will as determined byvariations in the position of the operators accelerator pedal, modifiesthe automatic ratio 5 selecting mechanism, and for all ratio changes,

likewise establishes the action of the driving coupling which is to takedrive in the new ratio. The invention utilizes fluid pressure means inthe example given herewith, for interconnecting the ele- 0 ments of thecontrol system, and the actuation system, whereby the described effectsare accomplished.

In the application of the invention due consideration has been given toautomatically controlled fluid pressure servo actuated systems of theprior art having superficial similarities. Such showings of step ratiogearing having fluid pressure actuation of selected friction members,and some form of governor or equivalent selection control are ordinarilybased on the general idea of having the automatic action duplicate asnearly as possible, the manual control functions of a vehicle operator.It follows therefore, that a method of automatic freewheeling which insome way may release drive on one or more of the driving elements, forexample, through by-pass or release of the fluid pressure, provides acommonly understood means for establishing a neutral dwell during theshift interval. The adaptation of the so-called automatic clutch controlto assist in automatic gearshifting is a well-known expedient. Itscontrol by the engine accelerator pedal is also known in the prior art.

Since portions of the present invention bear a' superficial resemblanceto devices of the above noted character, it is believed of primaryimportance to distinguish clearly from them, wherein'the utility andtrue purpose of the structures in my invention will be apparent, andwherein its unique and novel features will be understood.

A very considerable simplification in the number of control pedals,handles and buttons is desirable for safety reasons in present day motorvehicles.

The interconnecting means of my invention, linking elements moved by agovernor with other elements moved'by the engine accelerator pedalprovide a fully automatic speed ratio selection control for a form ofstep-ratio transmission in which there is no established neutral dwellperiod during the shift interval. Instead, there is a distinct overlapduring the interval, arranged through the proper selection of operatingprinciples, so that the new driving element begins to assume drive at agiven torque capacity value, while the element which is being releasedceases to drive only when that value is established. The doubling up ofthe controls for these functions eliminates a number of members which anordinary gearbox control presents for driver manipulation.

The invention describes a primary method of regulating the shifttransition torque capacity values, discloses a secondary method whereinthe ratio setting itself varies the values, and discloses a furthermethod of control which evaluates speed and torque indices, for acompletely automatic ratio shift sequence.

This is in no sense a freewheeling control of a friction clutch or groupof clutches, but is a control for a step-ratio transmission havinginherent torque overlap during the shift interval, whereby theconditions of drive and the intent of the driver are reflected in themanner in which the transition between fixed ratio steps isaccomplished.

My invention therefore combines the two features of driver participationin ratio selection, and the control over the torque capacity of thedriving-elements, at values always above a predetermined minimum torquevalue, the coordinating means being fluid pressure connected.

One outstanding advantage of the system of the invention is theelimination of torque shock between theengine and the load. Ofequivalent importance is the fact that it eliminates the veryconsiderable momentum and power loss common to step-ratio changingdevices which do require the neutral dwell, or no-drive period for theshift interval. It will be readily understood that each time an enginedecelerates to idling accompanied by the slip loss of frictionengagement of the new ratio, the ener y expenditure to bring therotating mass of the engine to a low relative speed, and to raise itagain to a new level, is expended without return. There is a loss offuel energy during the idling period, and the further slippage loss ofthe friction elements assuming drive. With my system, the inertia lossis only the small difference between the inertia value at one enginespeed and the value at a second engine speed proportional to the stepspeed ratio difference between adjacent speed ratios. My inventioneliminates the idling loss, and shows a negligible slip loss. These arefacts well demonstrated in actual practice, and are not in any sensetheoretical, having been demonstrated against the performance of theother type of devices in which a neutral dwell shift is required.

The performance of a number of types of infinitely variabletransmissions, automatically controlled, is known, and affords acomparison guide for my invention against the ideal, as well as for myinvention against those forms utilizing the neutral-dwell principle.This discloses that my provision of the continuous, overlapping torqueactuation method, with a constant control available for the overlappingvalue as well as the driving capacity values, in conjunction withdriver-dominated automatic ratio controls which select ratio accordingto the traffic conditions and the ability of the vehicle engine, yieldsprofoundly new and different results. As far as the vehicle driver isconcerned, the free-will ability to drive with full acceleration, orwith advantageous ratio for economy, gives a new activity of responsenot found in the compared form using the neutral-dwell principle, aswell as a new degree of smoothness and effortless operation.

The invention further provides calibration means wherein the variouscontrol factors may be adjusted to accommodate variations in the type ofoperation demanded of the vehicle in which the installation is placed.

It is believed that the demonstration of the present inventionestablishes an unquestionably close approximation to the resultsobtained with continuously variable transmissions of the friction type,wherein an ideal selection pattern for speed ratio yields full automaticcontrols for both economy and performance ratios.

Among other objects, the invention discloses a form of governor whichprovides unique means for controlling sequences of automatic ratio shiftthrough variations in fluid pressure, and solves a problem ofcoordinating the speed effect with the efi'ect of driver will on acontinuously operative cycle.

Modifications of the form of governor and associated ratio selectormechanism are given to provide those skilled in the art with sufllcientdata to adapt the method herein described to a wide range of vehicle andpower installation applications.

Further improvements are likewise shown in the interlocking of theovercoming manual controls with the automatically operating controlssuch that while a regime of full automatic speed ratio shift ispermitted, the manual controls may overcome and compel shift to aselected ratio for emergency purposes.

Additional objects are achieved in the interconnection of meteringvalving, automatically operated, through fluid pressure columns with theselector valving determining ratio, wherein shift interval torque may bevaried according to the driving conditions. Herein, likewise, areprovided modifications of the method, for a fully useful disclosure ofbroad applicability.

An additional object is the providing of a special form of ratio controlfor variable speed gearing in which during the automatic shift period asimultaneous shift of ratio determining elements is required, thecontrol method involving an overlapping of the net torque from oneelement to another, wherein a maximum of smoothness and minimum of sliploss occurs during the transition.

A general object is to approximate in efficiency with a step ratiogearing, that of the ideal infinitely variable gear, through particularmethods of selecting the ratio and regulation of the torquecharacteristics for the ratio transition period.

These and other objects will appear in the subjoined text andaccompanying drawings.

Figure 1 is a vertical, longitudinal section of the transmissionassembly of my example, with the engine shaft at the right.

Figure 2 is a transverse vertical section at 2-2 of Figure 1, showingthe servo pump assembly at the left and one form of governor at theright.

Figure 3 is a transverse section at 3-3 of Figure 1, the detail of thegears and clutch parts being omitted, the showing being directed to theactuation system. Figure 3a is similar to Figure 3 showing the parts ofthe actuator system for the front unit, at approximately 3a--3a ofFigure 1.

Figure 4 is a schematic diagram including the pump and governor ofFigure 2, the actuation system of Figures 1, 3, and 3a, and theauxiliary controls for coordinating the responses of the automatic andthe manually moved elements providing a full automatic shift among thefour available forward ratios. Figure 4a is a diagram of a camplatemoved by the operator-controlled linkage of Figures 2 and 13 forcontrolling the lever 320 and valve I68 of Figure 4.

Figure 5 is a sectional view of a modified form of governor to that ofFigure 2.

Figures 6 and 7 represent valve construction porting and details for anassembly of parts as shown in Figure 8.

Figure 8 is a schematic diagram including the servo pump of Figure 2,the two-stage hydraulic governor of Figure 5, the actuation system ofFigures 3 and 3a, the compensator control of Figure 4 and the valvearrangements of Figures 6 and 7. Figure 8 also shows the pump regulatingvalve and the lubrication system connection.

Figure 9 is a. typical diagram of four forward speed ratios with upshiftand downshift points given as examples of an application of my inventionto a motor vehicle, wherein the arrows pointing downward refer to speedsof engine and output shafts for increasing transmission ratio, and thearrows pointing upward refer to downshift changes at the noted relativespeeds.

Figure 10 is a typical pressure response curve of the hydraulic governorof Figure 2. the marked points referring to the ratio controlcharacteristics of- Figure 9. Figures 9 and 10 studied together, showthe coordination requirements for a typical shift pattern according tomy invention, wherein the objectives noted preceding are obtained.

Figure 11 is a diagram similar to that of Figure 9, wherein the ratioshift points are taken at somewhat different relative s eeds for thedesired forward ratio changes. from those of Figure 9. These arecoordinated to the two-stage pressure governor of Figure 5, thecharacteristic pressure curves of which, with respect to the controlsystem of Figure 8, and the valve arrangement of Figures 6 and 7, are asshown in Figure 12. It will be noted that the curves G and G arecombined to yield the resultant curve (3.

Figure 13 describes the lever system controlled by the operatorshandlever for shifting the forward-reverse gearing of Figures 1 and 2,and the cam plate of Figure 4a, or the cam plate 2I2' for valve I68 ofFigure 8.

In Figure 1, the general arrangement of the driving parts of thetransmission assembly are shown with the engine-connected shaft 5 at theright carrying integral gear element I. Between shaft 5 and the engineany standard form of clutch may be used. Shaft 5 is supported in the endwall of casing 2 at bearing 6, and in turn supports intermediate shaft 8in pilot sleeve 69.

Shaft 8 terminates at the left in annulus drum II, the hollow portionpiloting shaft 2| in bushing I3. Splines 9 of shaft 8 mount sliding gearI9 which may mesh with the internal teeth I' of gear I, or with reverseidler gear I8.

Countershaft I5 is rigidly fixed between the endwall of casing 2 and web4, and supports countershaft cluster gear body 28 on bushings I4, whichreceive oil from helical cuts 35 through drilled passages to bediscussed later. Teeth I6 of body 28 constantly mesh with gear I so thatbody 28 runs at a fixed ratio to engine speed at all times. Teeth II ofbody 28 constantly mesh with idler gear I8. When gear I9 meshes withteeth I of gear I, shaft 8 runs at engine speed; when I9 meshes withidler I8, shaft 8 runs reversely to shaft 5.

Gear I9 is moved along splines 9 by fork I88 attached to or integralwith slider I84 of rod I8I fixed between web 4 and the end wall ofcasing 2. Shifter shaft I83 of Figure 2 extends in arm I02 to cam headI86 moving in the slides of shifter I84. attached to rocker arm I89 aswill be described in detail later.

Gears I13 and I14 are fixed to body 28 and shaft 8 respectively, anddrive the fuel servo pump assemblies to be described, the cross-shaft II6 carrying gear I15 meshing with gear "4.

The transmission output shaft 58 mounted in casing 2 at bearing 49supports the rear end of shaft 2I, its front end being supported asdescribed preceding. Thrust bearings I8 and 41 locate shaft 2I withrespect to shafts 8 and 58, respectively. and thrust bearing I8transmits thrusts between shaft 8 and shaft 5.

Gear teeth I2 of drum II deliver input'torque to planets 24 rotating onspindles 23 supported in carrier 22 rotating with shaft 2I. Sun gear 25also meshing with planets 24 is integral wth sleeve 26 and drum 28rotatably supported on shaft 2| through the sleeve of carrier 22. andreaction forces are taken through web 29 of drum 28, rotating withsleeve 26.

Extension 34 of carrier 22 is attached thereto by rivets 3 I, and issplined at 48 to support clutch plates 33 which mate with clutch plates36 arranged to rotate with drum 28 by intersecting bolts 64. Internalseats for springs 88 are formed on drum 28; presser plate I4 rotatingwith drum 28, and serving to sustain extension forces from springs 88.

Portion 29 of drum 28 is recessed at II to form cylinders for clutchpistons I2 and piston pins Shaft I83 is rocked by external linkage I3arranged to coact with plate I4. Drilled passage I9 admits fluidpressure to cylinders 1| through porting to be described later. Thestepratio gearing herewith described is illustrative, and may bereplaced by other forms, wherein the principle of having a predeterminedminimum torque capacity available at all times is utilized, as will bedescribed later.

When fluid pressure is exerted on pistons 12, springs 88 are compressed,plates 33-36 are loaded for a predetermined torque capacity, and thecouple thus established requires gears I2 2425 to rotate as a unit withthe associated parts above described. When fluid pressure is releasedfrom cylinders II, springs 88 expand, and drive is removed from plates33-36. Brake 88 surrounding drum 28 may then lock the drum againstrotation, and require carrier 22 and planet gears 24 to planetate aroundthe nonrotating sun gear 25, through servo and control means to bedescribed.

Two net speed ratios are therefore available between shafts 8 and 2|according to whether clutch 33-36 is driving or brake 88 is applied todrum 28.

The above describes the structure and operation of the so-called frontunit of my example. Shaft 2I, as the output shaft of the front unit, isthe input shaft for the rear unit. Sun gears 31 and 38 are integral withor attached to shaft 2|, and mesh with planet gears 43-44 respec tively.Clutch hub 59 splined to rotate with shaft 2I is externally splined at61 to mount clutch plates 68 corresponding to clutch plates 33 of thefront unit.

A composite drum 39 bearing mounted on the sleeve of hub 59, providesbraking surface for brake 98, and is further supported at bearing 66 onshaft 50. The drum 39 is attached to annulus gear 42, the web of whichsupports springs 89. Presser plate I8 mounted to rotate with drum 39 isbored to receive piston pin 11 which transmits force from piston I6 incylinder I5, to plates 68 and to plates 55 arranged to rotate with drum39. The clutch plates 55-68 are identical with plates 36-33 of the frontunit; and are loaded by fluid pressure in drilled passage I9 similarlyto the actuation method for the front unit.

Annulus gear 42 meshes with planet gears 44.

Carrier 54 for planets 44 is attached to drum 52 having internal gearteeth 5I meshing with planet gears 43. Carrier 46 of planets 43 isintegral with or attached to output shaft 58. When' brake 98 is lockedto drum 39, annulus gear 42 is held against rotation, and planetation ofcarrier 54 and planets 44 occurs, a predetermined speed component beingapplied to annulus 5|. Simultaneously sun gear 31 is applying torque tocarrier 46, and instead of a speed ratio derived solely from the gearingcouple of 3I-435I, the net speed ratio of shaft 58 to shaft 2I is theresultant of the compounding of the motion of annulus 5| and sun gear 31through planets 43 upon carrier 46. In the present instance this is areduced forward speed ratio. 4 When clutch 55-68 is loaded by fluidpressure in cylinder I5 and brake 98 released, the couple thusestablished between shaft 2| and annulus 42 causes the assembly torotate as a unit.

The above description of the so-called rear unit provides two net speedratios, direct drive, and geared drive, so that with the front unitcombination, the overall number of forward speeds available is four aswill appear following.

In combining the preceding mechanism, it should be understood that Ihave associated two fundamentally different forms of planetary gearingsin series for the deliberate purpose of obtaining the maximum number ofproperly distributed speed ratios. This is not possible with ordinarycombinations of identical forms of gearings in that, for example, shouldthere be two units such as my front unit serially connected betweenpower and load, there would be only three net forward speeds, sincetheir reduction ratios would be identical. The actuation means andmethod herein described may take other forms, it being only essentialthat loading of friction elements engaged or sustained by fluid pressureand unloading of the friction elements constituting the torque carryingor sustaining means be likewise fluid pressure operated.

I, therefore, by utilizing dissimilar planetary forms in two-speedseries arrangement, obtain a proper distribution of speed ratiosavailable over an approximate uniform range, not otherwise available incompounded series gearing of this type.

Speed ratio of rear unit.

"Speed ratio of front unit.

In a series gear arrangement of this type the product of the individualreduction ratios is equal to the overall reduction through both units.

In the view of Figure 2, the forward and reverse shifter arm I09 rocksshaft I03, arm I02 and cam head I06 in slides I of shifter I04 whichslides on rod IOI, moving gear I9 as described. Gear I15 driven by gearI14 of Figure 1 drives shaft I16 and pump rotor I12, the meshing statorgear I83 of pump assembly I99, driven by movement of the vehicle.Countershaft body 20 and gear I13 drive gear I11, hollow shaft I18 andpump rotor I1I, with which mating stator gear I82, the primary group ofthe pump assembly I99 is found. The latter is subject to engine speedrotation.

The stator gears I83 and I82 are spindled in the housing, and sumpsuction pipes I9I and I10 feed to suction space I95 for both units. Thepressure outlets of the double pump assembly are shown schematically inFigure 4. The normal directions of rotation of both rotors I12 and I 1|is such as to add the resultant pump pressures, but when drive is inreverse, it will be seen from inspection of Figure 2 that rotor I12 willrun backward, therefore the lower capacity for this unit as comparedwith the I1I--I82 unit. In this event, the subtraction, or loss ofpressure is not sufficient to prevent positive pressure output from thepump assemblies.

Small gear I15 which drives shaft I16 also "drives governor shaft IIOpinned to the flange I II of the hydraulic governor shown at the rightin Figure 2. Flange III is attached to governor body II2, which hascounterweight portion I I3, and is carried at the outer end in gland H4.

The gear I82 (Fig. 2) is driven by meshing gear I1I of hollow shaft gearI11. The latter meshes with gear I18 fixed to countershaft body 20constantly geared through I61 to input shaft 5 which runs at enginespeed. This carries the speed index of the engine to gear I82.

Gear I83 is driven by meshing pump gear I12 and shaft I I 0, whichlatter receives drive through gears I15, I14 and shaft 8. Since rotationof shaft 50 is carried back to shaft 8 through the two units, in whichthere is never an interruption of torque, the pump unit I12-I83 alwaysyields pressure, even if I91' be disconnected, whenever shaft 50 hasrotation.

Valve II5 moves in the bore sections IIB of body II2, its portings H1,H8, H9 and I20 being connected through drilled passages I2I, I22 tooutlet pipe I23 receiving oil from pump manifold 200 of Figure 4 and topipe 220 which delivers modified pressures to the servo control units ofFigure 4.

When forward drive is established by meshing of gear I9 with gear teeth1' of Figure l, the governor partakes of engine speed, by means of theintervening gearing described. The governor may be said to be ahydraulic one in that it delivers a governed pressure according tochanges in speed.

In detail, the governor device of Figures 2 and 4 consists of body H2 inwhich two externally connected passages I22 and I24 are drilled, thefirst being joined by piping I23 to the pressure main 200 of the servoand lubrication pump I99 of Figure 2, the second to governor output line220.

The body I I2 is radially bored at I I8 to accommodate valve member II5,which is normally held by spring I25 retained by slotted screw cap I26.

The bore ports in order radially from the rotation center of the casing,are the inlet port II9, the servo port II8, the exhaust port I20, andthe control port II1. Inlet port II9 connects to oil pump lead I22.Servo port II8 of bore II6 has two outlets, one leading to the governoroutput line I24, and the other consisting of drilled passage I2I leadingto control port III. The exhaust port I20 dumps overpressure from excesslift of valve I I5 through the outlet shown.

Valve H5 at its outer end is formed into stop boss I28 adjacent largediameter boss II5b. Below 51) is smaller diameter boss II5a, the innerend of the valve II5 being formed into a small diameter boss II whichfits the smaller part of bore II6. See Figures 2 and 4.

The hydraulic governor casing consists of a cast body II 2 attached toflange III, pinned to governor shaft IIO of Figure 2. The outer sleeveend of the casing H2 is drilled radially at I 3I and I32 extending fluidpressure leads I22 and I24 to match with passages I33 and I34 in glandII4 which connect to pipes I23 and 220 respectively.

The body is formed into a counterweight I I3 to balance the mass of therotating assembly. Sealing rings I 35 prevent linkage of oil from glandI I4, at the bore.

The hydraulic governor as a controlling pressure regulator operatesthrough the interaction of pressures on the faces of bosses H511 andII5b of valve H5, and through the centrifugal force on valve II5 workingagainst spring I25, these forces being arranged by design to yield a netoutput pressure curve such as line O'A of Figure 10. This is anapproximate square-law curve, in which the modifying forces establish aseries of consecutive pressures which may constitute an approximatestraight line function, if such be desired. Heretofore, in speed ratiocontrol deproviding selective operation characteristics such as ailordedby my device.

The diagram in Figure is of especial value in focusing attention on theunique characteristics of my hydraulic governor. By providing a scale ofdistributed pressures at which the various speed ratio shift functionsmay take place, the equilibrating of the servo control devices to thoseselected pressures for the given functions is simple', and exact withincommercial requirements.

For example, the following table of selected pressures delivered by thepump represents the predetermined shift characteristics desired,,foradvanced or retarded throttle:

Shift efl'ects re uired 5? At eng. Throttle q ou S R. P. M. settingpressure Front unit Rear unit 6 200 (Downshilt below 200 R. P. M.) 8 680Bet Downshiit. 13 1150 Rat Upshiit. l5. 6 1330 Rel: Upshift.

37 2420 (Prevention manual downshiit 3rd or 4th to 2nd or 3rd) 49 2800Adv ps t 58 3100 Adv Upshiit. 3700 Upshilt.

The review of the disclosure in Figure 4 now to be described will makeclear how the above table is applied.

Figure 3 is a cross-section view of the actuation arrangement of thebrake and brake cylinder assembly for the rear unit. In this figure,brake drum 39 may be gripped by band 98 anchored at fitting 9I inhousing 2, by adjustable stud 92 and nut 92a. The movable end 93 of bandcarries pivot 98a for thrust rod I98 fitting notch I92 of rocker I93pivoted at I94 and actuated by piston rod 298.

Brake cylinder 292 is attached to, or integral with casing 2, and hasbolted spring retainer cup 293.

The details of the elements inside of cylinder 292 are now given. Piston29I slides in the cyl-- inder, and is attached to rod 298. Springs 91,91a, 91b, press the piston upwards, tending to apply brake 90.Compensator piston 294 slides in the smaller section 295 of the cylinder292, and rides on the rod 290. Compensator pressure line 218 connects tocylinder section 295, and servo pressure line 219 leads to cylinder 292so as to act upon piston 29I alone. Abutment 214 held by spring 910 andsliding on a guide extension of spring retainer 293 may be struck by pin291 attached to piston 294, but free to move in a drilled passage inpiston 29I. Pressure in 295 may be exerted on piston 294, compensatingthe action of the resistance of the springs 91, 91a and 91b on linepressure in 219, as will be described further. The actuation systemdescribed herewith is especially adapted to provide an overlappingtorque effect during change of ratio, so that a constant minimum torquevalue is maintained between input and output shafts for reasons to bedescribed in detail later in this specification. The spring arrangementsprovide stored energy utilized in both control of loading and regulationof the rates of loading and unloading as will be seen.

Figure 3a is a cross section view of the brake and brake actuatingassembly for the front unit. Brake drum 28 may be gripped by band 80anchored at 8I in casing 2 by stud 82. The movable end 83 of band 80carries pivot 80a for thrust rod I88 fitting notch I82 ofrocker I83pivoted at I84, and actuated by piston rod 288.

Brake cylinder 282 is attached to, or integral with casing 2, and hasbolted spring retainer cup 288.

Piston 28I slides in cylinder 282, and on rod 280, and is engageable byabutment 286 fixed to the rod, and by springs 81 and 81a, which tend toapply brake 80 of the front unit through the above linkage. An abutment288 sliding on a guide extension of retainer 283, is held by spring 81b,and may strike abutment 288.

Compensator piston 285 slides in the cylinder section 285' of 282 and isattached to rod 280. Spring 284 may transfer a given force from downwardmotion of 280 and 285 to piston 28I. Pressure from compensator line 211may be exerted in 285, compensating the action of the resistance of thesprings 81 and 81a on line pressure in line 218 connected to cylinder282, and arranged to remove the load of springs 81 and 81a. whileapplying the clutch 33--36 of Figure 1 as will be described later.

Having described the various group units, and how they operate withinthemselves, this text will now trace the interconnections among thegroups, including the coordinated action of the control elements bywhich predetermined variations in torque capacity of the drivingelements are obtained during the shift intervals, for maximum smoothnessand avoidance of shock effects.

Figure 13 gives a conventional linkage of handlever I5I attached toshaft I53 rotating arm I54 to reciprocate link I55 to rock lever I09 forshifting the gearing for forward and reverse, and to rock shaft 3I0b andcamplate 2I2 or its equivalent 2 I 2', through lever 3 I 0a.

The pump being driven as indicated in Figure 2 produces line' pressurein manifold 200.which is delivered to line 238232 to the rear unitcontrol valve port 324, line 238-233 to the input port 243 of thedifferential valve 240, line I23 to the input port II9 of the hydraulicgovernor, and through line 234 to the input port 288 of the actuatingvalve I50 for the front unit.

Whenever the engine or the vehicle, or both are in motion, pump manifold200 delivers fluid pressure to all of these points.

The hydraulic governor outlet pressure in Figure 4 is carried in line220 to both control cylinders 300 and 400 for both units, and throughconnection 222 to port 2I8 of plunger 2I5.

Referring to Figure 4a, valve I88 of Figure 4 is rocked by bell-crank320 between positions registering ports 324 and 325 or ports 325 and 328as previously described.

When the camplate 2I2 of Figure 4a is in either of the reverse or lowpositions, pin 2 in slot 2I3 is rocked counterclockwise, forcing valveI68 to occupy the down position.

When the camplate 2I2 is in neutral, the valve I88 by reason of thecontouring of slot 2I3, is compelled to shift clockwise, putting thevalve in the up position, or such that pump pressure may flow tocylinder 292 of the rear unit and to clutch line 19'.

In order to enable the driver to shift from direct drive in the rearunit, to 2nd speed, that is, to lock the brake 90 on drum 39, rod 2 andplunger 2I5 in bore 2I8 held down by spring 2", abuts lever 320, forexerting an upward thrust when fluid pressure from the governor throughline 222 is capable of overcoming spring 2 I 1, aided by fluid pressurein port 2I8, delivered from line 212', connected to the servo cylinderline of the rear unit, and acting on plunger 2I9.

This can only take place, however, at pressures corresponding togovernor speeds of approximately 2000 R. P. M., so that the operatorcannot abuse the engine by making a manual downshift to 2nd speed, whenthe higher governor and engine speed ranges are in force.

Connections 220a and 220b join line 220 to cylinders 300 and 400respectively.

Line 212 carries pressure from port 325 of valve I68 to the clutchcylinders 15 of the rear unit, and also to cylinder 292 and to piston29I, as in Figure 3.

Differential valve 240 meters pressure received from pump passage 238 tocompensator line 211, joined to piping 211 to compensator cylinders 285'and 295 of both units. Line 211 is also connected to servo controlcylinders 400 and 300 through connections 228 and 229 respectively.

Line 218 leading from port 261 of valve I50, delivers fluid pressure toclutch cylinders H and pistons 12 of the front unit, as well as to brakecylinder 282.

The servo control cylinder 300 for the rear unit is ported at 430 withconnected piping 43I therefrom to port 432 of servo control cylinder 400for the front unit, through pipe extension 43I, and line 212-43I joinsboth ports 430 and 432 to actuator cylinder 292 for the rear unit.

Figure 4 shows the pedal connected rod 36I attached to lever I43 throughan adjustable clevis. Lever I43 is mounted on shaft I44 rotatable in thecover plate (not shown) bolted to casing 2, and inside the cover'plate,the lever I45 is mounted to swing with shaft I44 as' the rod 36I isreciprocated.

Lever I45 is formed with a cam end, and at its opposite end is formedwith eyelet I21 arranged to intersect stop pin I29 which limits thetravel of the assembly of throttle pedal connected parts thus fardescribed.

The extension 32I of valve 240 locates metering spring 322 which pressesplunger sleeve 323 intersecting the cam end of lever I45. Sleeve 323 isarranged to transfer force to valve 240 entirely through spring 322 forpart of its travel upward, and thereafter to act upon extension 32I ofvalve 240 directly.

The diagram of parts in Figure 4 describes the complete servo controland servo actuation system. Attention is directed to the fact thattheservo control units act as pressure regulators for the actuatorunits.

For convenience in following the paths of fluid pressure through thevarious channels, the element groups will first be defined. The doublepump assembly, designated by numeral I90, is as described in Figure 2,the servo system outlet casing or manifold 200 delivering pump pressureto the hydraulic governor valve II5, the differential valve 240, thefront unit valve I50, the rear unit valve I68, and from these valves thecontrolled pressures are delivered to the servo control and actuationmembers.

The assembly of the servo control cylinder 300 regulates the operationof valve I68 which controls the actuation of the rear unit of Figure 1through pressures in clutch cylinders 15 and brake cylinder 292. Theassembly of the servo control cylinder 400 governs the operation ofvalve I50 which controls the actuation of the front unit of Figure 1through pressures in clutch cylinders 1| and brake cylinder 282.

Force applied to toggle arm 304 by rod 303, positions servo valve I68 ineither of two positions; admission of fluid pressure from line 238 tothe brake servo cylinder 292 and clutch cylinders 15, at port 325 shownin Figure 4, or release of pressure from port 325 and the cylinders15-292 to exhaust, through port 328.

The servo control arrangement for the front unit consists of cylinder400 having two bore sections 40I and 402, in which pistons 403 and 404may slide. Piston 403 is attached or integral with control rod 405; andpiston 404 is integral with secondary piston 406, sliding insub-cylinder 401. Spring 409 normally acts to hold piston 403 to theleft, and spring 4 acts on sliding abutment M0 to hold it flush withshoulder 4I2 of cylinder 400.

The throttle connected valve control of Figure 4 consists of a valvemember 240 occupying bore 24I in-valve housing 2I0, having oil pumpconnected port 243, reaction port 242, output connected port 244 andexhaust porting 245. This assembly comprises the throttle pedal meansfor establishing the operation of the speed ratio control devices ofFigures 3 and 3a.

Figure 4 provides also a schematic representation of the servo controlon the valving for both the front and rear units, controlling the servoactuation of the clutches and brakes of the said units.

Servo control cylinder 300 for the rear unit contains piston 30I workingin bore 302 adjacent bore 305. The piston 30I is rigidly attached tocontrol rod 303 pivoted at 306 to toggle arm 304. Spring 301 normallypresses piston 30I to the right of bore 302, and spring 309 pressessliding abutment 3I0 against shoulder 3 of cylinder 300. At the right,web 308 of cylinder 300 sets off a second compartment 3I2, in whichpiston 3I3 abuts rod 303a.

Toggle arms 303 and 304 pivoted to the casing 2I0 at 3I5, act togetherthrough attached snap spring 3I6. Toggle arm 303 is pinned to link 3I1,which in turn is pivoted to valve rocker arm 320 at 3I9. Rocker 320,pivoted to the casing 2I0 at 32I is attached at 32I to the master valveI68 for the rear unit, sliding in bore 330, ported at 324, 325 and 326.Port 325 is connected to the actuator cylinders 15 for clutch 55-60 ofthe rear unit and cylinder 292 for the operation of the brake 90 ofFigure 3. Port 324 connects through pipe 238 to pump manifold 200. Port326 is for exhaust. Valve I68 in the up position joinsports 325-324supplying pump pressure to actuate clutch 55-60 of the rear unit, andtake off brake 90 of that unit. In the down" position, valve I68relieves the pressure causing clutch 55-60 to disengage, permittingbrake 90 to be applied, sealing off pump port 324.

With advance of the engine throttle pedal I 4|, rod 36I rocks lever I43,shaft I44 and lever I45 counterclockwise, delivering an increasingstress to spring 322. This applies an upward pressure on valve 240tendingto close off fluid in port 243 from port 244. As the pedal isretracted, the port opening is increased, permitting greater flowbetween the pump line 238 and compensator line 211.

The hydraulic governor assembly of Figures 2 and 4 provides a novelmethod for obtaining a wide range of control functions to meet therequirements for motor car speed ratio control.

The radial movement of valve I I5 from the shaft center is conditionedby centrifugal force of rotation of shaft IIO on the valve II! asopposed by spring I25, and by pressure conditions in passages I22 andI24. The valve is ported to achieve a metering pressure control effectproportional to speed of shaft II 0, as will be discussed following.Depending upon the dimensions and force in spring I25, the outputpressure curve may be made to approximate the square of the appliedspeed, or to assume correlative values to correspond to the shiftrequirements of the fiuid pressure control system departing somewhatfrom the square law, as is found expedient The exposed end area of thelower boss 5a to pump line pressure from manifold 200 is smaller thanthe exposed area of the upper boss 5b, therefore, when shaft speed islow and pump pressure just building up, the initial movement of valve H5is more rapid than in the immediately following stage when not only is alarger area exposed, but also the resistance of spring I25 is greater.

When port I I8 is open to line pressure from I22 through H9 and, bylifting of valve H5, the exposure of the top of the valve to pressurefrom passage I2I to port I I1 tends to hold the valve I I5 stable at agiven radial distance from the shaft center, while applied speed andpressure conditions remain Within predetermined limits.

At the lowest speed, 200 R. P. M., it is desirable that the pressurefrom the pump assembly diminish to a point below which both front andrear units will be in their reduction speed ratios. Assuming forwardacceleration of the vehicle by opening of the engine throttle, thegovernor output pressure will'rise to 13 pounds at 1150 R. P. M. asshown in the chart in Figure 10, when the piston 403 of Figure 4 willmove against spring 409. At light throttle, valve 240 will be permittingfull compensating pressure in passage 21'! so that piston 404 will notinterfere with move-- ment of rod 405, and toggle 4I54I6 will snap valveI50 to the right, admitting servo pressure to clutch cylinders II andbrake cylinder 202 of the front unit.

This will establish shift to "direct" in the front unit, which by thetable of ratios, is "2nd speed ratio.

At 2800 R. P. M., and 49 pounds pressure on the curve of Figure 10, thefull throttle action has moved valve 240 of Figure 4 to shut-oilposition so that there is no pressure acting to hold out piston 404.Line pressure in 220 must therefore be sufficient to overcome springs409 and M0 and move piston 404 out of the way. Spring 4 is so positionedthat it cannot move piston 403 far enough to the left to produce adownshift. At this pressure point the front unit will shift to its"direct or to 2nd speed ratio, even if the throttle pedal has been heldin-fully depressed position.

Between these two pressure-speed points, the variation of enginethrottle and engine speed interact so that up-shift may occur at aninfinite number of points between.

This relationship is also shown in the, graph of Figure 9.

After the shift to 2nd speed ratio has been made, the front unit is indirect; but the rear unit is required to shift up to direct, with ashift in the front unit to reduction gear," in order to obtain 3rd speedratio.

when the governor delivers sufficient pressure, such as at 1330 R. P.M., augmented by compensating pressure on 3I3, the net pressure acts onpiston 30I, overcomes springs 301-309 shifting 303 to the left, snappingtoggle 3030-304, putting the rear unit in direct drive, under lightthrottle conditions.

If engine speed after the above shift occurs is below the downshiftpoint (680 R. P. M.) for the front unit, that unit will be downshiftedby spring 409 overcoming the governor delivered pressure, giving 3rdspeed ratio.

But if engine speed after the shift is greater than 680 R. P. M., thetransmission is in 4th speed under light throttle conditions.

Full throttle shift from 2nd to 3rd speed is obtained at 3100 R. P. M.,when there is sufficient governor pressure to overcome springs 301 and309 without compensation pressure on 3I3.

When the rear unit shifts, the rear unit servo pressure acts on piston406 of the front unit servo control, sufficient to augment spring 409andproduce downshift of the front unit.

v However, the ratio will remain in 3rd at full throttle until agovernor speed of 3700 R. P. M. is reached, when governor pressure issufficient to overcome springs 409-4 and the rear unit pressure onpiston 400.

In the examination of the ratio shift chart of Figure 9 it must beappreciated that the straight lines radiating from theoriginrepresenting net speed ratios between transmission input andoutput shafts, proportioned in slope according to the gear diameters andcenter distances in the example. The slopes of the lines will varyaccording to selected speed ratios, and are not limited to the exactratios used herewith.

The ratio pattern is herein repeated for demonstration purposes:

Low -s.1'z 2nd 2.23 3rd. 1.42 Direct 1.00

the numbers referring net input speed to output speed of 1.

The low and 2nd lines on the chart encompass an area A between enginespeeds of 1150 to 2800 and 800 and 1960 respectively.

The shift pattern of Figure 9 is laid out to show engine speed on the OKscale and transmission output shaft speed on the OY axis. The

four lines radiating from O and marked respectively low, 2nd, 3rd anddirect" represent the availablespeed ratios given as examples in theforegoing description. I

The diagram of Figure 10 represents a typical set of controlcharacteristics afforded by my hydraulic governor arrangement. Thevertical scale represents net output pressures provided in passage I24leading from the governor of Figure 4 to the servo control system of thefigure; and the horizontal scale represents net engine speeds. Thecurved line beginning at 200 R. 'P. M., engine speed and 6 poundspressure, and engine at 3700 R. P. M. and 80 pounds pressure, is arepresentative pressure curve derived from the hydraulic governor of myinvention.

The form of hydraulic governor shown in Figure 5 consists of a rotatablevalve body '2' attached to shaft 0' driven by gear I15, and spindled ingland 4' which has a running fit with the rightward extension of thebody I I2.

Two valves 5' and III' occupy bores H0 and I I3 respectively, the boresbeing open to exhaust at their outer ends. Valve 5' has bosses I2I' andI22 in sequence from the centerline outward, the boss I2I" fitting awider portion of thebore 6'. Boss I22 terminates in weight I23, to bereferred to as W in this specification.

Valve III has bosses I25 and I26 in sequence outward from thecenterline, the boss I25 fitting a wider portion of bore I I3. Theextension of boss I26 constitutes a weight member I24, to be referred toas W in this specification.

A central vent port I20 for bosses I2I and I25 is drilled into valve 2'at I21, the lead opening to the sump 23I or the suction side of the pumpassembly I99.

The lowermost port II1 connects by passage III! to port H9 open to theupper face of boss H5 and to the lower face of boss I2I, both portsconnecting in turn to passage I24 and annular ring I28, cut on theextension of body H2.

The uppermost port I29 connects with port I30 by passage I3I, the portI30 being exposed to the upper face of boss I25 and the lower face ofboss I20, both ports I29 and I30 being connected by passage I32 toannulus I33 cut in the extension of casing II2.

Pressure ports I34 and I35 for each of the valves III and H5 arecross-connected by passage I36 leading to annulus I31 cut in theextension of easing II2.

Annulus I28 is in register with line I38; annulus I33 is in registerwith line I39 and annulus I31 with line I40, the lines leading fromfixed gland II4 to the servo and control elements of Figures 6 and 7.

It will be observed that W is a larger weight than W and by thewell-known centrifugal effect, valve III upon a given increase inrotation will first tend to move radially from the position shown.

Assuming a constant pump pressure in line I40 and ports I34 and I35, theopening of port I35 is caused by centrifugal force shifting the valve III so that port I3I intersects the lower edge of boss I26, and permitsmetered pressure to be exerted on the upper face of boss I25 and on thelower face of boss I26, and because of the dissimilarity in the areas ofthe bosses, the net metered pressure acting together with thecentrifugal effect tends to reach an equilibrium point corresponding toa given speed of rotation of body II2.

At this equilibrium position, the lower face of boss I 26 tends torestrict the flow from port I35 to port I30 by a given amount, so thatthe net pressure existing in passage I32 and line I39, to supply a givendegree of control, such as G will depend primarily upon the factor ofgovernor body rotational speed, as modified by the differential areapressure factors of bosses I25 and I26.

The curve O-I of Figure 12 shows a series of pressures for givengovernor speeds, corresponding to the effect of valve and weight actionas provided by the above noted interacting pressure and centrifugalforce mechanism.

The centrifugal and pressure effect noted above is also utilized in theoperation of valve II5. As centrifugal force on valve H5 and weight I23increases, the opening of port I34 by passage of the upper edge of bossI22, permits pump line pressure to be exerted on the lower face of bossl2I and on the upper face of boss I22, and because of the dissimilarareas of the bosses, the metercd pressure acting together with thecentrifugal effect, tends to reach an equilibrium point corresponding toa given speed of rotation of body 2'.

" stage G of the governor.

At this position, the upper face of boss I22 tends to restrict the flowfrom port I34 to port II9 by a given amount, so that the net pressureexisting in passage I24 and line I38, to supply a given requirement,such as G", will depend upon the factor of governor speed, as modifiedby the differential area pressure factors of bosses I2I and'I22. Thecurve -11 of Figure 12 shows a series of pressures developed by valve Hin the output pressure passages connected thereto.

It is assumed that normally constant leakages are experienced in thepassages beyond lines I33 and I39 wherein the speed effects of thegovernor are to be utilized. Excessive pressure caused by blocking ofeither of the lines is automatically taken care of by the bosses I2I andI25, which being of larger areas than bosses I22 and I26, force thevalves inward toward the centerline until the external exhaust stems HIand I42 relieve the excess, whereupon the designed pressure for a givenspeed of body II2 will be resumed.

Fluctuations in pump pressure in line I40 are taken care of by valve 350of Figure 4. If not so handled, the differential-area pressure tendingto move valve II5 against centrifugal force, for example, remains thesame, since the metering by boss I22 of the line pressure from passageI36 and port I34 allows the valve to feel output or metered pressureonly. A loss of pump pressure would only determine the maximum operatingpressure of the governor at a lower level.

In Figure 8 the pump pressure regulating valve 350, in a bore of easing2, receives pump pressure in ports 352 and 353 from pump pressure outlet354, and delivers uniform line pressure to leads 332 and 238 throughcheck valve 355, and port overpressure port 351 dumps back to the sumpat high pump speeds, as controlled by spring 358, whose tension isadjustable by nut 359. Port 368 and line 369 feed the lubrication inlet21 of Figure 1 through passage H in shaft 8, and a similar passage inshaft 2I. At given pressure conditions, lines 213, 216, 238 and 332 maydrain back to sump through line 369, when the valve 350 is in theposition shown.

In deriving curves 0-I and O-II of Figure 11, the net pressure effectupon the output system moved thereby must be considered along with 0 theareas and derived pressures of valves H5 and III, in that a parallelcombining of forces on the output system can be made to yield aconsiderable family of resultant curves, typified by OIII--II of Figure11. This will be clearer when the dissimilar areas of plungers 53I--533and 50I--503 are considered.

The two control valve groups in Figures 6 and 7 are displaced, as shownin Figure 8 in the easing 500, the group to the left being for thecontrol of the rear unit, and that to the right for the front unit.

The uppermost plunger 50I at the right is open to port 502 connected toline I39 of Figure 8, receiving pressure from the primary pressure stageG of the hydraulic governor of that figure.

The adjacent abutting plunger 503 is open to port 504 connected to lineI38 of Figure 8. and receives pressure from the secondary pressure Thebore 505 is open to exhaust at 506.

The front unit control valve 5I0 sliding in bore 509 of casing 500 hasan upper boss 501 and a lower boss 508. Spacer pin 5 in drilled hole 5I2transmits the force of plungers EM and 503 moss-1a open to port 5I5connected to line 218 of Figure 8, joined to cylinder 282 and to line 19of the clutch actuation system of the front unit.

Intermediate the ports H3 and M5 is exhaust port 5I6 open to sumpdrainage. The lower end of boss 588 is open to exhaust at 5", and abutspin 5I8 sliding in drilling 5I9, which pin in turn transmits effort toplunger 528, open at its upper end to port 52I connected to compensatorline 216' of Figure 8, and at its lower end to exhaust at 522.

' Plunger 528 abuts plunger 523 whose upper face is open to exhaust at522, and whose lower end is open to port 524 connected to pump line 213of Figure 8.

Spacer pin 525 sliding in drilling 526 bears on the lowermost plunger538 whose upper face is open to exhaust port 521, and whose lower faceis exposed to port 528 connected to line 213', rear unit servo, ofFigure 8.

The valve group at the left of Figure 6 will now be described. Theuppermost plunger 53I is open to port 532 connected to line I39 ofFigure 8, receiving pressure from the primary pressure stage G of thegovernor.

The adjacent abutting plunger 533 is open to port 534 connected to lineI38 of Figure 8 and receives pressure from the secondary pressure stageof G of the governor. The bore 535 is open to exhaust at 536. The rearunit control valve 548 sliding in bore 531 of casing 588, has upper boss538, lower boss 539, and skirt 54I. Spacer pin 542 in drilled hole 543transmits the forces of plungers 53I and 533 to valve 548, abutting boss538, which is open to port 544 connecting to rear unit servo line 212 ofFigure 8.

The interspace between bosses 538 and 539 is shown as open to port 545connecting to pump line 213 of Figure 8 joined to pump main 238; andopen to port 546 connected to rear unit servo line 212 of Figure '1,joined to cylinder 292 and to line 19' of the clutch actuation system ofthe rear unit. The lower face of boss 539 is open to port 541 connectedto pump line 213 of Figure 8. Bore 548 encloses spring 558, whichpresses control valve 548 upward, with a given force, for reasons to bedescribed later. Intermediate port 549 between ports 544 and 546 is opento exhaust.

The lower end of skirt 54I of control valve 548 engages spacer pin 55Isliding in drilling 552 and casing 588. The pin projects into port 553engaging the upper end of plunger 554, the port being connected to line212 of Figure 8. The bore 555 of plunger 554 is open to exhaust at space556, and engages spacer pin 551 sliding in drilling 558, and pressing onplunger 568, whose upper face is exposed to port 559 connected tocompensator line 216 of Figure 8 for the purpose of delivering throttlecontrolled pressure to plunger 568.

The lower face of plunger 568 is exposed to port 56I connected to pumppressure line 213 of Figure 8. Spacer pin 562 connects plunger 568mechanically with the lowermost plunger 565,

the upper face of which is open to exhaust port 564, and the lower faceof which is open to port 563 connected to line 34I of Figure 8, the linebeing controlled by the movement of manual valve I68, as shown'in Figure8.

The schematic diagram of Figure 8 describes the essential controls forthe two stage governor ratio selection'system combined with theregulatory means for predetermining a clutch control in each of thefront and rear units.

The governor assembly is shown in outline in the upper left portion ofthe drawing, the gland II4 having three connecting lines, the first ofwhich, I48, is connected to the pump main 236; the second of which, I39,delivers primary pressure output from the governor; and the third ofwhich, I 38', delivers secondary pressure output from the governor.

The details of the construction of the governor are shown separately inFigure 5.

The pump I99 schematically described in the lower left portion of thedrawing, as has been explained, continuously operates as long as eitherof the engine or the output shaft of the transmission assembly are beingdriven. The suction line I98 connects to the sump of the transmissionand the outlet pressure main 238 delivers to lines 238, 332, and 213.Line 213, as will be-described, delivers pump pressure for, servoactuation for both the front and the rear units of the assembly to frontunit control valve 5I8 and rear unit.

control valve 548. Line 332 supplies pressure to compensator valve 248;which in turn provides compensation pressure for both units, so that theclutch capacities thereof may be controlled in accordance with motion ofaccelerator pedal. Line I48 obtains pressure from main 238 throughmanual control valve I68.

The regulation of the torque capacities ofthe friction elements so thatthe driving mechanism will be prepared for operator-determined changesin torque demand, and also for changes in the driving conditions, is anessential feature of my invention. The metered output pressure of valve248 accomplishes this result.

Compensator valve 248 shown schematically in the lower center portion ofthe drawing has been described in detail in Figure 4. Line 332 from thepump main 238, is opened to port 264, its pressure being permitted topass through bore 339' to port 335, and thence to lines 216 and 216'leading to the compensation system.

The manual control valve I68 shown in diagram in the left center portionof the drawing receives pump pressure from main'line 238 delivered toport 263 of bore 339. The adjacent port 262 delivers pressure to thegovernor inlet I48 and the lowermost port 348 may deliver pressure toline 34I' leading to port 563 of the rear unit control valve systemdescribed in detail in Figure 6. The control valving for the rear unitshown in the approximate center of the drawing is also described indetail in Figure 6. Ported passages 545, 541, and 56I are joined to line213 connected to pump main 238; and ported passages 544, 546, and 533,are connected to line 212 leading to the rear unit cylinder 292 and toline 19' connected to gland 281, which supplies fluid pressure to theclutch cylinder 15 of the rear unit shown in Figure 3. Side connection213 from line 213 is connected to port 528 of the front unit valveassembly of Figure 7.

Compensator porting 559 of the valve control for the rear unit isconnected to line 216 leading to port 335 of compensator valve 248; andlikewise to line 216' leading to port 52I of the valve assembly for thefront unit control shown in detail in Figure 6.

Ports 5 and 524 are connected to line 213 connected to the pump pressuremain 238. Likewise ported passages 5'I3 and. 5I5 connect to the servopressure line 218 for the front unit cylinder. Passage 218 leads togland 281 of Figure 1 and through porting I9 to the cylinders 'II forthe clutch pistons of the front unit, shown in Figure 3.

In the following discussion of the operation of the preceding mechanism,for convenience, the pressures from the primary stage of the governor inline I39 will be referred to as G and from the secondary stage as G. Thesummation of these will be noted as G.

The output pressures from valve 240 in lines 211 of Figure 4 will bereferred to as C, since the compensation pressures coacting with thevalving controls are of identical nature. Line pressure from the pumpI99 in main 238, and acting in various leads connected thereto, will bereferred to as P.

In operation, the control system of Figures 6 and 7 in conjunction withthe governor of Figure 5 and the throttle connected valve 240 of Figure4 is arranged to operate all shifts through the four forward speedsautomatically, by the combination of governor developed pressure withaccelerator pedal controlled pressure.

It will be assumed that pump I99 is operating and valve I55 is in itsuppermost position connecting ports 252 and 283, so that the hydraulicgovernor begins to provide pressure in line I38 and line I38. However,the W and W weight characteristic creates a pressure on plungers 53I and"I according to line 0-111 of chart of Figure' 12.

On the chart line O-II represents the governor characteristic of valveIII weighted at I 25 line 0-11 represents the governor characteristic ofvalve 5' weighted at I23; and line 0I1III provides the summation ofthese two lines because the pressures in lines I38 and I39 are added bythe plungers I-503 for the front .unit and by the plungers 53I533 fortherear unit, as will be apparent.

The force of G and G on plungers 50I and 302 is endeavoring to shiftcontrol valve 5I0 to admit pump pressure from port 5 to port 5I5.

With a light throttle, valve 240 is delivering a compensation pressureto ports 52I such that it aids the governor pressure in overcoming pumpline pressure applied to plunger 523, which is normally tending to holdvalve 5.! in the up position, or the front unit in low."

When the combined G and G pressure reaches 20 pounds, for example, orthe governor shaft II! a speed of 750 R. P. M., the additionalcompensation pressure, for example C, with G will overcome P; shiftingvalve 5I0 down, to produce a shift to direct in the front unit.

However, at full throttle, the compensation pressure C will reduce tonearly zero, requiring G to rise to a value of about 63 poundscorresponding to a governor R. P. M. of 2600, for example.

These values are obtained byproper proportioning of the various valveand plunger areas as will be apparent to one skilled in the art.

To prevent hunting of valve 5I0, port 5I3 is connected to servo line andport 5I5, so as to add a sustaining pressure tending to hold the valvedown, once having been moved to direct drive piosition (in this case 2ndspeed) by the eflfort o G.

The shift from 2nd speed to 3rd speed ratio is accomplished by movementof the valve 5, ac-

cording to G and (3 pressures, opposed by the effect of spring 550 andline pressure acting in port 5" on boss 539 and on the base of sleeve 5.

A variable throttle motion delivers a variable pressure to port 559 fromvalve 240 of Figure 7, and plunger 580 may move against the pressure inport "I connected to the pump line 213. This sets up two conditions;first, at light throttle, full C pressure may cancel the effect ofpressure in I, on 560, so that the G value opposing spring 550 and linepressure in 541 may shift valve 540 "up; at 46 pounds governor pressure,or 1150 R. P. M. of the governor, which will shift overall speed ratiofrom 2nd to 4th.

At full throttle, when driving in 2nd, the C pressure in 559 vanishes,and at a pump pressure of approximately 73 pounds, or a speed of 3300 R.P. M., G1 and G are able to overcome the spring 580 and line pressure in541, and in "I, and engagement of the clutch of the rear unit begins tobe established.

This extra load on the engine, due to the application of a mechanicaladvantage of the load against the engine by the upshift, brings down theengine speed.

The engine at the instantaneous load will then drop in speed to below2400 R. P. M., for example.

A new control condition is established. Now, the values of G and C havefallen off to below pounds, which by examination of the chart of Figures11 and 12 places full line pressure on plunger 523, and pressure fromline 213 to plunger 53!, which since 52I at full throttle is drained,causes the force of G and G to be overcome; whereupon valve 5I0 shiftsup, and the front unit is put into reduction gear.

Now with the rear unit in direct and the front unit in low, the overallspeed ratio is 3rd speed.

For the 3rd to 4th shift at light throttle, the G values acting on thefront unit valve assembly are aided by throttle controlled pressure inMI, as noted previously, but the resisting force to the upshift tendencyis increased by the pressure in port 528 acting on plunger 530, so thatthe governor has to go to a higher speed such as 1025 R. P. M., or apressure of 37 pounds, to shift valve 5I5 down, or to the direct driveposition for the front unit, which will, of course, establish overall4th speed ratio.

At full throttle, pressure in port 52I vanishes, so that the linepressure value in 524 under plunger 523, plus the additional pressure onplunger 530 from port 528 opposes G directly, so that the shift to 4thwill not occur until a governor speed of 3670 R. P. M. or 79.5 poundspressure is reached.

It should be noted that when driving in 4th speed ratio, the pressure in5I3 is assisting the governor in holding the valve 5I0 in the down"position. To produce a downshift, with the accelerator pedal fullydepressed, it will be noted that the C pressure in port 52I vanishes,and that the force exerted by the sum of 523 and 523 is sufficient toovercome the G value and also the line pressure in port "3, provided theG valve does not exceed 60 pounds, or the governor a speed of 2400 R. P.M.

When, however, the pressure in MI is increased as with light throttlesetting, the elements involved in downshift are: plunger 523 feeling theline pressure in 524; plunger 53!! feeling rear unit servo pressure in528; resisted by the C pressure in port 52I on plunger 520; linepressure in port 5| 3 acting on boss I" and the G values. not ex- Iceeding 18 pounds, or corresponding to a governor speed of 700 R. PQM. V

. When the values on 623 and 636 overcome the values on 526, 501, 503,and 501, the valve Billis shifted up, putting the front unit inreduction gear; or establishing overall 3rd speed ratio.

When in 3rd speed ratio, the front unit is in reduction gear and therear unit is in direct. It

pressure in port 559 on plunger 554 opposes.

The area of plunger 554 is slightly smaller than the area of 566,yielding a small residual pressure tending to shift valve 540 "up at ahigher G value when the C value is not active. However, when a 0 valueis present, the residual pressure is cancelled, and the downshift willoccur at a lower G value, such as three pounds, or about 300 R. P. M..

If when driving in 2nd speed it is desirable to shift to low at fullthrottle, this may be accomplished up to 800 R. P. M. of the governor,or at below 22.5 pounds; in which the pressure in port 524 on 523overcomes the pressure in port 5I3 and the G value. At light throttle,the C value in 52l opposes the pressure in port 524 on 523, so that theresidual force on 523 is suirlcient to overcome pressure in port M3 anda G value of 1.2 pounds or at a governor speed of 150 R. P. M.

It' should be understood that the speed and pressure ranges between fullthrottle and light throttle control conditions constitute scales ofpressure and speed over which an infinite number of incremental shiftpoints may occur, as torque demand determined by the operator inmovement of the accelerator pedal is equilibrated with the pressureconditions of the systemyand that the values used in the above exampleare limiting orend points between which the shift effects occur.

These values are purely arbitrary, and are so chosen to illustrate theapplication of the principles of the invention. They may be modified bythose skilled in the art without departing from the principles ofinvention disclosed herein.

In Figure 8 manually controlled valve I68 in bore 339 has 3 positions:first, when connecting ports 262-463, joining pump main 238 withgovernor input line I40; second, when connecting 262263340' so that line3" carries pump pressure to port- 563 for plunger 565. The lowermostposition cuts off pump port 262, dumping both lines I40 and 3M so thatno automatic shift can take place when driving in reverse. These threepositions correlated with the handlever control of Figure 13 correspondrespectively to "neutral-high, low, and reverse, as is apparent from thecamplate relationship of Figure 8.

In the preceding descriptions the valves which control directly theshift of ratio' in the units have been considered as ratio selectorvalves. For example, in Figure 4 valve I68 controls the change ofratioin the rear unit, and valve I50 the change in the front unit. InFigures 6, 7 and 8,

the valve 540 controls the rear unit,and 5H1 con-,

trols the front .unit. In Figure 8, valve I68 is the master controlvalve which directs the pump pressure to the governor in accordance withthe movement of the operators handlever of Figure 13.

It should be further borne in mind that the sequences of the fullautomatic ratio control functions are arranged in a predeterminedpattern wherein at a low governor speed, the throttle effect is notavailable to select ratio while the governor is determining upshift; atmiddle speed ranges of the governor, the throttlecontrol ef fect coactswith that of the governor; and at high governor speeds the throttleeffect is again inhibited. These three control ranges or regimes arereadily apparent from study of the charts of Figures 9 to 12 inclusive,in conjunction with the foregoing text. It will be noted that amongother advantages, this method protects the engine from abuse when thevehicle momentum is low and also when it is high, leaving the middlerange for full activity of the driver-will control component.

Attention is again directed to the coordinated means for regulating thetorque capacity of the friction drive elements, and the continuousminimum torque, or torque overlap characteristic of the servo actuationsystem associated therewith. The complete combination including fullautomatic shift through governor controlled elements coasting withdriver-will controlled elements through fiuid pressure governing andgoverned means is believed unique with my invention, and provides forthe first time a step ratio gearing controlled in such a way that theresults approximate those achieved by automatic infinitely variabletransmission systems. That this approximation is factual, isdemonstrated by the charts of Figures 9, 10,11, and 12 derived fromactual experience with the constructions of my invention. I

The present demonstration includes inventions of my prior applications,insofar as parts of the whole maybe compared therewith, but representsnew unitary combinations as outlined in the foregoing specification.

A wide application ofv the principles of my invention is possible,wherever a variable speed transmission is required to yield speed ratiochanges between a variable speed power source and a variable torqueload, and wherein automatically controlled regimes of selectiveperformance and economy are to be utilized to meet changing driveconditions. The field of application includes motor driven trucks,buses, railcars, military tanks and vehicle equipment of all types, andalthough the specific demonstration herewith is for a drive of thepassenger car type, I reserve the right of application of the principlesof the demonstration to these other forms, as outlined in the scope ofthe appended claims describing the spirit of the invention.

I claim: Y

1. In variable speed gearing controls, in combination, an engine, anengine throttle control, a power shaft driven by said engine, a loadshaft, ,a variable speed transmission connected to said shafts, servomotor means adapted to shift the speed ratio of said transmission, avariable speed pump driven by said shafts including a suction inlet, apressure outlet, hydraulic pressure means controlling the effective pumpoutlet pressure within a predetermined range of pressures, said meansbeing also subject to spring pressure, speed ratio selecting means, agovernor responsive to centrifugal speed connected to said outlet and tosaid selecting means. and means movable with said throttle controladapted to vary the effect of said governor upon said selecting meansaccording to the position of said throttle control. 7

2. In automatically operated power transmission mechanisms, incombination, a power shaft, a load shaft, a pump driven by one of saidshafts, a governor operating at the speed of one of the shafts,including a rotating valve, a pressure inlet space connected to saidpump, a pressure outlet space, a variable speed transmission joiningsaid shafts operated by fluid pressure actuating means connected to saidpump, coacting speed ratio selecting devices connected to said outletspace adapted to control the setting of said actuating means accordingto the speed of said governor, and resisting means operating to opposethe action of said governor over a speed range whereby .speed ratio isautomatically selected according to values of predetermined speed ofsaid governor and said resisting means.

3. In variable speed mechanisms, a fluid pressure source, a pressuremain connected to said source, a speed ratio selecting device connectedto said pressure main, a hydraulic speed-responsive governor adapted toprovide a predetermined range of variable pressures between said mainand said device, including a rotatable valve, and modifying meanscoacting with said device and operator-operable whereby the net effectof speed response of said governor upon said device is varied accordingto the will of the operator.

4. In power control devices, a variable speed transmission embodyingspeed ratio actuating means, a hydraulic governor including a primarycentrifugal valve and a secondary centrifugal valve, pressure outletsconnected to said valves, speed ratio selecting valving connected tosaid actuating means, and to said outlets, and manual means adapted tovary the action of said governor upon said valving according to the willof the operator.

5. In power transmission mechanisms, a first variable speed transmissionunit, a second variable speed unit driven thereby, a ratio selectingdevice for said first unit, a ratio selecting device for said secondunit, a hydraulic governor arranged to deliver variable pressures tosaid devices automatically according to speed values impressed upon it,and interconnecting means establishing coaction between said deviceseffective to prevent a ratio shift in one unit simultaneously with aratio shift in the other unit.

6. In automatic power controls, for motor vehicles, in combination, anengine controlled by a throttle pedal, a power shaft driven by saidengine, a variable speed transmission driven by said shaft and driving aload shaft, actuation means for said transmission, ratio selectingdevices operative upon said actuation means, a hydraulic governorarranged to deliver variable pressures to said devices according tospeed values of one of said shafts, and variable pressure means moved bysaid pedal coacting with said governor and effective upon said deviceswhereby the driving ratio of said transmission is selected and actuatedaccording to the operator torque demand and the driving conditions.

'7. In automatic controls for variable speed gearing, in combination, agovernor including a centrifugal valve arranged to provide a primaryrange of predetermined pressures with variable speed applied to thegovernor, a second centrifugal valve arranged to provide a dissimilarsecondary range of predetermined pressures with variable speed appliedto the governor, and pressure responsive means adapted to respond to thecombined pressures of both said ranges.

8. In automatic controls for variable speed gearing, in combination, ahydraulic governor embodying a presure inlet, and a pressure outlet, apressure resisting means, and a centrifugal valve acting against saidmeans arranged to position itself according to applied speed and to thepressure of said outlet.

9. In pressure responsive controls for variable speed gearing, incombination, a centrifugal governor embodying a resisting means, andprovided with inlet and outlet pressure, and a speed responsive valvesubject to said inlet pressure, said resisting means, and to said outletpresure.

10. In pressure responsive controldevices for variable speed gearing, acentrifugal governor provided with inlet and outlet pressure spaces, andoutlet pressure control means comprising a primary valve of a givenpredetermined centrifugal force characteristic, a secondary valve of adissimilar predetermined centrifugal force characteristic, both saidvalves being subject to said inlet pressure while responding todissimilar outlet pressures.

11. In variable speed gearing, in combination, a power shaft, a firstvariable ratio transmission unit driven by said shaft, a second variableratio transmissionunit driven by said first named unit and connected toa load shaft, fluid pressure servo means arranged to actuate ratio shiftin said units, selecting devices acting upon said servo means, agovernor driven by one of said shafts including centrifugal valvemechanism connected to said devices, and additional control meanscoacting with said governor and said devices effective to modify theaction of said mechanism upon said devices whereby speed ratio changesare established in said units according to the combined effect ofgovernor speed and positioning of said control means.

12. In automatically controlled fluid pressure servo actuated systems ofchange speed gearing for motor vehicles, an engine, a driver operatedspeed control member for said engine, an output shaft, a step ratiogearing connecting said shaft to said engine including fluid pressureactuated elements for establishing various speed ratios therebetween; afluid pressure system comprising a source, actuators for said elementsand intervening control valving, members of said valving beingresponsive to variations in fluid pressure, a governor embodying a valveresponsive to centrifugal force arranged to deliver continuouslyvariable fluid pressure to said control valving according to changes inthe centrifugal force of said valve, and a second valve controlled bymovement of' said member for modifying the effect of saidgovernor-embodied valve on said control valving according to enginespeed determining positions of said pedal.

13. In speed ratio controls for power transmissions, a power shaft and aload shaft, variable speed gearing therebetween having friction elementsfor establishing selected speed ratios of drive through said gearing,said elements being fluid pressure actuated, control valving operativeto select actuation of said friction elements, pressure responsive meansarranged to shift said control valving, and a weighted valve rotating atthe speed of one of said shafts effective to establish a continuousseries of pressures upon said pressure responsive means according to therotating speed of said weighted valve.

